Optimized offset strip fin for use in contact heat exchangers

ABSTRACT

An offset strip fin for use in compact automotive heat exchangers is disclosed. The offset strip fin has multiple transverse rows of corrugations extending in the axial direction wherein the corrugations in adjacent rows overlap in order that the oil boundary layer is continually re-started. The fin dimensions have been optimized in order to achieve superior ratio of heat transfer to pressure drop along the axial direction. In one aspect, an compact concentric tube heat exchanger has an offset strip fin located in an annular fluid flow passageway located between a pair of concentric tubes. The preferred range of lanced lengths is determined to be between 0.035&#34; to 0.075&#34; for periodically developed flow. Maintaining the lanced length in the regime of periodically developed flow is advantageous in that it gives a higher heat transfer coefficient than is achievable with fully developed flow. This also provides the added advantage that variations in the shape of the flow passages from the rectangular do not impact negatively on the heat transfer.

FIELD OF THE INVENTION

The present invention relates to offset strip fins used in compact tubeheat exchangers for use in automotive applications.

BACKGROUND OF THE INVENTION

Typical transmission and transaxle oil coolers employ tubular heatexchangers mounted in the outlet tank of the vehicle radiator. Theseheat exchangers include a cylindrical outer tube, an inner tube and aturbulizer placed in an annular passageway between the inner and outertubes. Oil is admitted to the annular passageway via an inlet portlocated at one end of the tube whereupon it passes through theturbulizer and is cooled and exits via an outlet port located near theother end of the tube.

Conventional turbulizers (also referred to as turbulators) which havebeen used in tubular heat exchangers typically consist of sinusoidalconvolutions or rectangular corrugations extending in rows axially alongthe length of the tubular heat exchanger. Adjacent rows in the flow oraxial direction are displaced from one another by half a convolutionthereby creating transverse rows of transversely aligned parallel slitsor apertures. The function of this geometry is to create artificialturbulence since as the hot oil flows through the heat exchanger andimpinges against the leading edge of the corrugations, the resultingexcessive form drag splits the oil flow sideways as it advances to thenext row of corrugations. This artificial turbulence is on the one handdesirable in that it results in enhanced heat transfer characteristicsbut is deleterious on the other hand in that it produces a significantcontribution to the pressure drop along the axial length of the heatexchanger.

Current design trends in the automotive industry are towards morecompact and aerodynamically efficient designs in an effort to increasefuel efficiency and accommodate new accessories such as pollutioncontrol devices and the like. This has led to a need to reduce the sizeof the radiator tank and hence a more compact concentric oil cooler isrequired. It has been found that down-sizing concentric oil coolersemploying conventional turbulizers results in a substantial increase inthe pressure drop along the axial length of the cooler. This higherpressure drop can produce deleterious effects on the oil pump therebyreducing the oil circulation rate in the cooling system.

Attempts have been made to minimize the oil pressure drop in the flowdirection by eliminating the artificial turbulence. This is achieved bychanging the turbulizer orientation so that the corrugations aretransversely aligned in circumferential rows with apertures through thecorrugations opening in the axial or flow direction thereby formingfluid flow passageways. The resulting structure does not createsignificant artificial turbulence and therefore cannot strictly bereferred to as a turbulizer but is more appropriately termed a fin. Thefin is comprised of a plurality of these circumferential rows (alsoreferred to as strips) of corrugations which extend in the axialdirection of the tubular heat exchanger. The walls of the passagewaysare periodically interrupted along the axial or flow direction, andcorrugations in adjacent rows or strips have been overlapped by 50% inorder to provide a continual restarting of the fluid boundary layers inorder to achieve high heat transfer properties. Fins having a geometrywherein adjacent rows or strips of corrugations are offset from eachother are typically referred to as offset strip fins (OSF). In thiscontext, offset refers to the fact that adjacent transverse strips areoffset from each other by a certain amount such that the corrugations inthe adjacent rows overlap to produce the interrupted flow passageways.

Recent theoretical studies (Sparrow, E. M. et al., Transactions of theASME, February 1977, p.4; and Sparrow, E. M. et al. J. Heat MassTransfer, Vol. 22, p.1613) suggest that there is considerable potentialfor achieving increased heat transfer and lower pressure drop using theOSF with the appropriate fin dimensions.

SUMMARY OF THE INVENTION

The subject invention provides an offset strip fin having a geometry anddimensions in a range suitable to provide optimized heattransfer-to-pressure drop ratios when utilized in compact heatexchangers for cooling automotive based oils.

In one aspect of the invention, an offset strip fin for use in a heatexchanger includes a plurality of transverse rows of corrugations wherethe rows are adjacent and extend in an axial direction. The corrugationshave a flat top portion and a flat bottom portion where both the top andbottom portions have the same width. The corrugations have a height in apredetermined range and a width in a predetermined range, with thepredetermined range of height being greater than the predetermined rangeof width. The corrugations in adjacent rows overlap with the overlappingcorrugations defining periodically interrupted flow passageways in theaxial direction. The lanced length of the corrugations in the axialdirection is in a predetermined range.

In another aspect of the invention, a parallel plate heat exchanger isprovided which includes a generally rectangular metal container withparallel top and bottom plates, one side having an entrance port and anopposed side having an exit port, and wherein the direction between theentrance and exit ports defines a longitudinal flow direction. An offsetstrip fin is disposed between the top and bottom plates wherein the finis provided with a plurality of transverse rows of corrugations, therows being adjacent and extending in the longitudinal direction. Thecorrugations have flat top portions and flat bottom portions, the topand bottom portions of the corrugations having the same width, the topportions being in thermal contact with the top plate and the bottomportions being in thermal contact with the bottom plate. Thecorrugations have a height in a predetermined range and a width in apredetermined range with the predetermined range of height being greaterthan the predetermined range of width. The corrugations in adjacent rowsoverlap to form periodically interrupted flow passageways in thelongitudinal direction. The corrugations have a lanced length in thelongitudinal direction in a predetermined range.

In a further aspect of the invention a tubular heat exchanger isprovided having an inner tube disposed within an outer tube with thespace between the tubes defining a passageway extending along the axialdirection of the tubes. An inlet port in flow communication with thepassageway and an outlet port in flow communication with the passagewayand spaced from the inlet port is provided. An offset strip fin isdisposed in the passageway between the tubes wherein the fin is providedwith a plurality of transverse rows of corrugations, the rows beingadjacent and extending in the axial direction. The corrugations have asubstantially flat top portion and a substantially flat bottom portion,the top and bottom portions having the same width, with the top portionin thermal contact with the inner surface of the outer tube and thebottom portion in thermal contact with the outer surface of the innertube. The corrugations have a height in a predetermined range and awidth in a predetermined range with the range of height being greaterthan the range of width. The corrugations in adjacent rows overlap toform periodically interrupted flow passageways in the axial directionwith the lanced length of the corrugations in the axial direction beingin a predetermined range.

In still another aspect of the invention a tubular heat exchanger forcooling automotive transaxle and transmission oil includes an outer tubehaving an inner diameter, and an inner tube concentrically disposedwithin the outer tube and having an outer diameter less than the innerdiameter of the outer tube. The space between the tubes defines anannular flow passageway in the axial direction of the tubes and the endsof the tubes are sealed together around the circumference. The tubesdefine an inlet port in flow communication with the annular passagewayand an outlet port also in flow communication with the passageway, theoutlet port being spaced from the inlet port. Also, an offset strip finis circumferentially disposed in the annular passageway extendingaxially between the ends of the tubes, wherein the fin comprises aplurality of transverse rows of corrugations, the corrugations definingflow passageways in the axial direction. The corrugations have a heightsubstantially equal to the difference between the inner radius of theouter tube and the outer radius of the inner tube, the corrugationshaving a flat top portion in thermal contact with the inner surface ofthe outer tube and a flat bottom portion in thermal contact with theouter surface of the inner tube. The top and bottom portions of thecorrugations have the same width, the width of the corrugations being ina predetermined range. The top portions of adjacent corrugations areseparated by a distance greater than the width of the corrugations andthe bottom portions of adjacent corrugations are separated by a distancewhich is less than the width of the corrugations. Also, corrugationshave a lanced length in the axial direction in a predetermined range.

BRIEF DESCRIPTION OF THE DRAWINGS

Preferred and alternative embodiments of the invention will now bedescribed by way of example only, with reference to the accompanyingdrawings, in which:

FIG. 1 is a perspective view of a preferred embodiment of a concentricheat exchanger according to the present invention;

FIG. 2 is a sectional view of the heat exchanger of FIG. 1 taken alongthe lines 2--2;

FIG. 3 is a perspective view of a portion of a fin in the flat orunwrapped form;

FIG. 4 is a front view of a fin showing the relative orientations ofoverlapping corrugations in two adjacent rows;

FIG. 5 is a sectional view of the wrapped fin of FIG. 4 showing therelative orientations of overlapping corrugations in two adjacent rowswherein the wrapped fin exhibits regular flow passages;

FIG. 6 is an enlarged view of the fin of FIG. 5 showing the relativeorientations of overlapping corrugations in two adjacent rows or strips;

FIG. 7 illustrates a) developing hydrodynamic flow in an offset stripfin dimensioned so as to prevent reaching the fully developed flowcondition for the given fluid flow rates and fluid properties, and b)fully developed flow in a rectangular passageway;

FIG. 8 is a sectional view of a wrapped fin which is on the verge ofexhibiting crossover;

FIG. 9 is an enlarged view of the wrapped fin of FIG. 8 showing therelative orientations of overlapping corrugations in two adjacent rowsat the limit of exhibiting crossover for the relative fin dimensionsshown;

FIG. 10 is a sectional view of a wrapped fin exhibiting crossover;

FIG. 11 is an enlarged view of the wrapped fin of FIG. 10 showing therelative orientations of overlapping corrugations in two adjacent rowsexhibiting crossover for the relative fin dimensions shown;

FIG. 12 is a cross-sectional view of a fin exhibiting highly unevenlyspaced and irregularly shaped flow passages;

FIG. 13 is a three dimensional plot summarizing the heat transferstudies on concentric heat exchangers using the LPD fins of the presentinvention wherein a plurality of fins with corrugation widths in a rangeup to a maximum of W=0.050" and lanced lengths L in the range 0.010" to0.270" have been studied;

FIG. 14 is a three dimensional plot summarizing the pressure dropstudies on concentric heat exchangers using various embodiments of theLPD fins (with H=0.105") of the present invention with corrugationwidths varied in the range 0.026" to 0.050", and lanced lengths variedin the range 0.010" to 0.270";

FIG. 15 summarizes the performance data of FIGS. 13 and 14 and similardata for LPD fins with H=0.130", indicating the optimal ranges forcorrugation width and lanced length where *=oil flow of 3 GPM; FinH=0.105"; Δ=oil flow of 3 GPM; Fin H=0.13"; =oil flow of 0.79 GPM; FinH=0.105"; and +=oil flow of 0.79 GPM Fin H=0.13".

FIG. 16 compares the heat transfer and pressure drop characteristics fortwo coolers of identical volume employing conventional turbulizers ofdiffering convolutions per inch (cpi) with the heat transfer andpressure drop characteristic of a cooler with a lower volume and whichutilizes an LPD fin, where □=conventional turbulizer with 5 cpi, withcooler dimensions being 1.0" dia., length being 12.8" c/c;+=conventional turbulizer with 3 cpi, cooler dimensions being 1.0" dia.,length being 12.8" c/c; Δ=LPD, 0.75" dia., length being 12.8" c/c; andL=0.044", H=0.1" and W=0.03".

FIG. 17 is the same as FIG. 16 but with different cooler dimensions anddifferent LPD fin dimensions; where=conventional turbulizer, 5 cpi, withcooler dimensions being 1.25" dia., length being 12.8" c/c+=conventionalturbulizer, 3 cpi, with cooler dimensions being 1.25" dia., length being12.8" c/c; Δ=LPD, 1.0" dia., length being 12.8" c/c; and L =0.044",H=0.1" and W=0.03".

FIG. 18 is the same as FIG. 16 but again with different coolerdimensions and different LPD fin dimensions; where □=conventionalturbulizer, 5 cpi, with cooler dimensions being 1.5" dia., length being12.8" c/c; +=conventional turbulizer with 3 cpi, with cooler dimensionsbeing 1.5" dia., length being 12.8" c/c; Δ=LPD, 1.25" dia., length being12.8" c/c; and L=0.044", H=0.1" and W=0.035".

FIG. 19 is similar to FIG. 16 but with still different cooler dimensionsand different LPD fin dimensions; where □=conventional turbulizer with 5cpi, with cooler dimensions being 1.75" dia., length being 12.8" c/c;+=conventional turbulizer with 3 cpi, with cooler dimensions being 1.75"dia., length being 12.8" c/c; Δ=LPD, 1.5" dia., length being 12.8" c/c;and L=0.044", H=0.1" and W=0.038".

FIG. 20 compares the heat transfer and pressure drop characteristics fortwo concentric coolers both having the same volume but wherein oneutilizes a conventional turbulizer where □=5 cpi (1.0" dia., 12.8" c/c)and the other an LPD fin Δ=LPD (1.0" dia., length being 12.8" c/c; andL=0.044", H=0.1" and W=0.03" and the other an LPD fin;

FIG. 21 is a partial sectional view of an alternative embodiment of anLPD fin illustrating two adjacent rows of corrugations of a fin as theywould appear in the wrapped form wherein the fin exhibits less than 50%offset in the flat form;

FIG. 22 illustrates another embodiment of an LPD fin similar to FIG. 21exhibiting more than 50% offset in the flat form;

FIG. 23 is a perspective view of yet another alternative embodiment ofan LPD fin in which adjacent rows of corrugations are offset by aconstant amount in the axial direction; and

FIG. 24 illustrates a perspective view of a flat plate heat exchangerutilizing the LPD fin of the subject invention.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

The geometry of the preferred embodiment of the offset strip fin andconcentric heat exchanger of the subject invention will be describedfirst followed by a discussion of the preferred range of the various findimensions and the experimental results from which these dimensions havebeen deduced. Reference will be made to the Figures wherein likenumerals refers to like parts.

Referring first to FIG. 1, a concentric tube heat exchanger 30 embodyingthe subject invention includes an outer cylindrical tube 32, an innercylindrical tube 34, an oil inlet port 36 located adjacent one end oftube 32 and an oil outlet port 38 spaced from inlet port 36 and adjacentthe other end of tube 32.

FIG. 2 illustrates a cross sectional view of heat exchanger 30 takenalong lines 2--2 of FIG. 1 wherein the outer diameter of inner tube 34is sufficiently smaller than the inner diameter of outer tube 32 so thatwhen tube 34 is concentrically disposed within tube 32, an annularpassageway 40 is formed therebetween along the axial direction of thetubes. Heat exchanger 30 is provided with an offset strip fin 42 whichis circumferentially disposed within annular passageway 40 and extendsbetween inlet port 34 and outlet port 36. The ends of outer tube 32 andinner tube 36 are sealed together around the circumference of the tubeends at 35 thus sealing fin 42 therein, see FIG. 1.

For reasons which will become apparent later, fin 42, having dimensionsfalling within a prescribed range to be set out below, exhibits asignificantly reduced pressure drop over conventional turbulizers andother offset strip fins and hence is referred to by the inventor as alow pressure drop (LPD) fin.

FIG. 3 shows a perspective view of a portion of fin 42 in its flat formwhile FIG. 4 is a front view of same. The portion of fin 42 shown inFIG. 3 comprises a plurality of generally rectangular shapedcorrugations 44 disposed in transverse rows (or strips) shown at 46, 48,50, 52 and 54. A complete fin such as would be found in heat exchanger30 comprises a plurality of these rows extending in the axial directionwhen the fin is annularly disposed within passageway 40 as indicated bythe arrows in FIG. 3. Corrugations 44 include a top surface portion 56,side portions 58 and bottom portions 60. Note that side portions 58 maybe structurally referred to as fins and hence the overall structure isreferred to as a fin. Corrugations 44 define apertures or flowpassageways 62 opening in the axial direction. When a fluid such as oilis flowing through fin 42 it will periodically encounter leading edges64 associated with corrugations 44.

Referring again to FIG. 3, corrugations 44 are characterized by thefollowing dimensions; fin thickness T, corrugation or fin height H,corrugation width W and row width or lanced length L. The fin thicknessT corresponds to the fin wall thickness against which the fluidimpinges, or leading edge 64 as it flows axially through the rows ofcorrugations 44. Since all the corrugations have the same height, thefin height and the corrugation height are the same hence fin height andcorrugation height refer to the same dimension.

The fin height H corresponds to the difference in the inner radius ofouter tube 32 and the outer radius of inner tube 34 since top portion 56and bottom portion 60 are in thermal contact with the inner surface ofouter tube 32 and the outer surface of inner tube 34 respectively whenheat exchanger 30 is fully assembled. Note that thermal contact betweentop portions 56 and bottom portions 60 with the respective portions oftubes 32 and 34 may be achieved in several ways including directmechanical contact or by forming a metallurgical bond such as bybrazing, the details of which will be determined by the particularmaterial used in the construction of fin 42 and tubes 32 and 34.

The lanced length L, also referred to in the literature as the offsetlength, (the former will be used hereinafter to signify L in order toavoid confusion with the percent offset of the fin to be discussedbelow) is the length of sides 58 of corrugations 44 in the direction offluid flow through fin 42 (as indicated in FIG. 3).

The corrugation width W refers to the width of the top and bottomportions of corrugations 44. Several different arrangements can occurand must be specified. First, the fin may be characterized by top andbottom portions having widths which are equal and thus the width refersto the width of both top part 56 and bottom part 60. Alternatively, thetop and bottom portions could have different widths in which case bothmust be specified separately. In the present invention, top part 56 andbottom part 60 have the same width W.

The percent offset in the flat form refers to the offset in adjacentcorrugations along both the top and bottom parts of the fin and isusually expressed as a percent. In the context of the present invention,since the top and bottom portions of the corrugations have the samewidths, therefore in the flat form the offset refers to the offsetbetween both top parts 56 and bottom parts 60. When the widths of thetop and bottom portions are of unequal length, then the % offset must bespecified for both the top and bottom parts of the fin. The amount ofoffset between corrugations 44 in fin 42 illustrated in FIGS. 3 and 4 is50%, however, as will be discussed later the amount of this offset isnot critical and may be more or less than 50%. The portions of the topand bottom parts of corrugations in adjacent rows which share a commonboundary are joined at those positions, such as is shown at 63 in FIG.3.

Referring to FIGS. 5 and 6, when fin 42 is placed within annularpassageway 40, corrugations 44 become distorted from their originalrectangular shape in the flat form. Overlapping portions of corrugationsin adjacent rows form periodically interrupted fluid flow passages 65 inthe axial direction. Due to the differences in circumferences of theinner surface of tube 32 and the outer surface of tube 34, the spacingbetween adjacent top parts 56 of adjacent corrugations 44 increaseswhile the spacing between adjacent bottom portions 60 of corrugations 44decreases, see FIG. 6. Once fin 42 is placed within passageway 40,corrugations 44 adopt a generally trapezoidal shape. Therefore, adjacentfluid flow passageways through the overlapping corrugations will havedifferent shapes and cross-sectional area but will nevertheless beregular or periodic along the flow direction. This results in flow pathswith differing resistances to flow which can, depending on the magnitudeof the differences, lead to significant flow maldistribution and hencepoor heat transfer.

Results of Studies To Determine The Optimum Range of Fin Dimensions Fora 50% Offset Fin

The inventor has carried out extensive and comprehensive studies todetermine the preferred fin dimensions which give optimized heattransfer-to-pressure drop ratios for a wrapped fin wherein the flowpassages are not all the same size or shape, see FIG. 6. The results ofthese studies are summarized herein.

In order to minimize the pressure drop along the axial direction andmaximize heat transfer in the direction normal to the fluid flowdirection, it is necessary to provide a fin geometry which on the onehand gives laminar flow through the flow passageways and maintains athin oil boundary layer while also minimizing flow maldistribution. Inaddition, the fin will preferably have a high surface area to present tothis thin oil boundary layer for efficient heat transfer. The highsurface area is achieved by decreasing the cross-sectional dimensions ofthe flow passages in the direction in which heat is transferred from theoil to the fin, i.e. at right angles to the walls of the passageway.

Referring to FIG. 7a, the periodically interrupted passageway wallsprovide for better heat transfer by maintaining the developing boundarylayer thin through the continual restarting of the boundary layers,shown at 66. In order to eliminate excessive form drag which occurs whenthe oil or fluid front impinges onto the leading edges of corrugations44, the fin thickness T should be as thin as possible. For materialsfrom which fins are typically fabricated such as alloys of copper,aluminum, brass, various steels and related alloys, the preferredthickness T for the fin has been determined to fall in the range from0.002" to 0.004".

The regularity of the flow channels will be determined in large part bythe relative relationship between the corrugation width W and the finheight H (see FIG. 3). At one extreme, highly irregular and unevenlyspaced flow passages result when overlapping corrugations in adjacentrows cross over along the inner circumference. The attendant decrease inheat transfer performance in the presence of crossover is found to bequite significant. For the 50% offset strip fin it has been determinedthat in order to avoid crossover between corrugations in adjacent rowsof the fin wrapped in the annular passageway, the fin height H shouldpreferably be less than 0.130" while the corrugation width W shouldpreferably be less than 0.050".

Referring again to FIGS. 5 and 6, the fin illustrated therein ischaracterized by the regular flow passageways 65 since both H and W fallin the preferable ranges (note FIG. 6 is a scaled up representations ofthe fin). The fin of FIGS. 8 and 9 (scaled up) is on the verge ofexhibiting crossover while the fin illustrated in FIGS. 10 and 11(scaled up) clearly exhibits crossover, the fin having a height Hslightly larger than the recommended upper limit of 0.130". Acorrugation width W greater than 0.05" shows a tendency to cross over,thus this establishes the upper limit on the corrugation widths for finswith heights in the range 0.100" to 0.130".

FIG. 12 illustrates a sectional view of a cooler 110 exhibitingextremely unevenly spaced and irregular flow passages 112 arising when afin 114 is characterized by a corrugation widths W and height H whichfall outside the prescribed ranges.

While the above established upper limits on fin height, thickness andcorrugation width provide for fairly uniform flow distribution, in orderto maximize the heat transfer and minimize the core pressure dropbetween the ends of the heat exchanger, the relative ranges for thecorrugation width and lanced length must be determined.

Referring to FIG. 7b, it is well known that superior heat transfercoefficients are obtained in the entrance region 100 of rectangular flowpassages 102 since they are characterized by developing hydrodynamic andthermal boundary layers shown at 104. The development of the offsetstrip fin is an attempt to exploit this effect. The hydrodynamic entrylength may be approximated by 0.05*H_(d) *R_(e), where H_(d) is thehydraulic diameter and R_(e) is the Reynolds number. This means that theratio L/H_(d) *R_(e) should not exceed 0.05 for hydrodynamicallydeveloping flow to exist. The Nusselt number N.sub.μ is given by theexpression N.sub.μ =h*H_(d) /k

where h is the convective heat transfer coefficient and k is the thermalconductivity of the fluid.

Recent theoretical studies have shown that for interrupted flow passagessuch as those produced with the OSF, see FIG. 7a, another type of fullydeveloped flow, known as periodic flow, exists rather than purehydrodynamically and thermally developing flow for lanced lengths in thehydrodynamically developing regime. This type of flow is characterizedby velocity and temperature profiles which vary along each strip but areinvariant from strip to strip at the same axial stations from theleading edge of the strip or corrugation. The mean laminar Nusseltnumbers N.sub.μ for periodic fully developed flow are significantlyhigher (2 to 5 times depending on the lanced length) than thecorresponding Nusselt numbers for thermally and hydrodynamicallydeveloped flow. Periodic flow in a non-rectangular flow passageways maystill give higher heat transfer coefficients compared to rectangularpassageways with fully developed flow. This factor outweighs anydeleterious effects of slight flow maldistribution arising in thenon-rectangular flow passageways resulting when the fin is in thewrapped form.

That there will exist an optimum lanced length L for achieving bothmaximum heat transfer performance and a minimum pressure drop along theaxial or longitudinal length of the heat exchanger can be understood forthe following reasons. Maintaining the boundary layer thin results inbetter heat transfer due to a shorter heat conduction path. Thus, as thelanced length L is decreased the heat transfer coefficients willincrease in the flow passages due to the continually decreasing heatconduction path length. For a given R_(e), a reduction in L/H_(d)results in an increase in the Nusselt numbers N.sub.μ (and hence heattransfer coefficients h) for the periodic flow. However, the rate ofincrease in N.sub.μ decreases as L decreases further and approaches anasymptotic value. Thus, there is no significant advantage to be gainedby choosing L less than this minimum value since the heat transfercoefficient h has reached a limiting value. In fact, from the point ofview of pressure drop, reducing the lanced length further may have anegative impact on the pressure drop in the axial direction. Thedimensionless pressure drop K_(p) is given by

    K.sub.p =2ΔP/ρ*v.sup.2 *t

where v is the flow velocity, ρ is the fluid density and ΔP is thepressure drop between the ends of the heat exchanger and t is the lengthof the heat exchanger between the ends of the heat exchanger. It hasbeen observed that larger pressure drops are obtained at smaller L/R_(e)*H_(d). Therefore, increasing the number of interruptions over thelength of the heat exchanger results in an increase in pressure drop.Note that scarfed or bent edges of the corrugations as well as theirfinite thickness will also contribute to higher pressure drops, thus thefin fabrication technique may play a significant role in the overallpressure drop of the cooler.

The results of heat transfer studies to determine the preferable rangesfor the corrugation width W and lanced length L will be graphicallydisplayed by plotting Nusselt number N.sub.μ versus L and W.

The results of pressure drop studies for the same range of corrugationwidth and lanced length will be graphically displayed by plotting thedimensionless pressure drop ΔP versus L and W.

FIG. 13 summarizes the results of heat transfer studies for a fin ofheight H=0.105" while FIG. 14 summarizes the corresponding pressure dropstudies. It is clear that over the entire range of dimensionlesspressure drop, K_(p), the peak for heat transfer generally occurs in therange of lanced length from 0.035" to 0.075" and corrugation widthmaintained in the range 0.030" to 0.050".

FIG. 15 graphically summarizes the data contained in the plots of FIG.13 and 14 wherein the ratios of Nusselt numbers (hence heat transfercoefficients) to dimensionless pressure drop are plotted against theratios of the lanced length to corrugation width for two different flowrates, 0.79 gpm and 3.0 gpm. FIG. 15 also summarizes data (not shown)similar to that displayed in FIGS. 13 and 14 but for a fin of heightH=0.130" at the flow rates of 0.79 and 3.0 gpm. Therefore the optimumrange for L has been determined for the reduced or downsized heatexchanger application.

The optimal ranges for the fin dimensions for a 50% OSF based on theabove results of fluid properties, fin structure, heat transfer, andpressure drop studies are summarized in Table I below.

                  TABLE I                                                         ______________________________________                                                      RANGE                                                           PARAMETER          FROM       TO                                              ______________________________________                                        Fin height H       0.100"     0.130"                                          Lanced length L    0.035"     0.075"                                          Corrugation width W                                                                              0.027"     0.050"                                          Fin thickness T    0.002"     0.004"                                          ______________________________________                                    

Referring now to FIGS. 16-20, the heat transfer and pressure dropcharacteristics for the concentric tube heat exchanger utilizing the LPDfin of the present invention are plotted and compared to those forconcentric heat exchangers employing conventional turbulizers. Fromthese plots it is clear that the former exhibit heat transfercharacteristics comparable to the later while exhibiting significantlylower pressure drops. Considering the differences in volume betweencoolers using the conventional turbulizers and those using the LPD finsin FIGS. 16 to 20, it is clear that the latter also exhibit comparableor better heat transfer performance-to-heat exchanger volume ratios thanthe former. The full advantage of the LPD fin of the present inventionover the conventional turbulizer is clearly demonstrated in FIG. 20where the dimensions of both heat exchangers are identical.

In light of the foregoing, a compact or downsized concentric heatexchanger utilizing an OSF fin has been disclosed which exhibits apressure drop significantly lower than that observed with concentriccoolers using conventional turbulizers. In addition, the heat transfercharacteristics of the former are comparable to or better than those ofthe latter. This improvement in the operating characteristics of thedownsized heat exchanger has been achieved by:

1) designing a fin with the appropriate fin height to corrugation widthto decrease the cross-sectional area of the fluid flow passagewaysnormal to the walls of the passageways in order to achieve both shortheat conduction paths normal to the direction of fluid flow and toprovide a large contact surface area between the passageway walls andthe fluid flowing therethrough; while simultaneously

2) maintaining the corrugation width to fin height ratio in theappropriate range to ensure the regularity of the flow passage profilein order to reduce flow maldistribution; and

3) determining the preferable ranges for corrugation width and lancedlength which result in hydrodynamically and thermally developingperiodic flow in the flow passageways at a reduced pressure dropcompared to that observed with conventional turbulizers.

It will be appreciated that the determination of the preferable rangesfor the lanced length L and the corrugation width W to produce the LPDfin was carried out on an offset strip fin with 50% offset with finheights in the range between 0.100" to 0.130" and corrugation widthsless than 0.050". As discussed above, OSF's with offsets greater or lessthan 50% will also be acceptable as long as the ranges of fin height Hand corrugation width W are such that regular flow channels areachieved. Specifically, as long as the dimensions H and W are such thatwhen the fin is disposed within the annular passageway no crossoveroccurs, deviations from 50% overlap are acceptable. FIG. 21 illustratesa blowup of a partial sectional view of a wrapped fin 120 characterizedby an offset less than 50% while FIG. 22 shows a partial wrapped fin 130with an offset greater than 50%. In both cases regular flow passages 122and 132 ar achieved in the wrapped form. Thus while the preferableranges for L and W for an OSF with greater or less than 50% offset arenot specifically disclosed herein, it will be understood that theinventor considers as part of the scope of the subject invention allcompact heat exchangers employing fins with offsets in the vicinity of50% which have been optimized with respect to the pressure drop and heattransfer to produce the LPD fin.

FIG. 23 illustrates another alternative embodiment of the fin of thesubject invention comprising an offset strip fin 150 with a constantoffset Q between the edges of corrugations 44' in adjacent rows. Theconstraint on the dimension Q will be that no crossover occurs when fin150 is in the wrapped form.

As mentioned above, the finite fin thickness and the presence of anyscarfing or bent edges will result in generally higher pressure drops.Thus it is desirable to have the thinnest fin possible.

While the optimized LPD fin dimensions have been determined for aconcentric tubular heat exchanger for automotive applications, it willbe readily apparent to those skilled in the art that the LPD findisclosed herein may be readily adapted for use in other heat exchangergeometries. FIG. 24 shows a flat parallel plate heat exchanger at 170comprising two plates 172 and 174 provided with an LPD fin 176sandwiched therebetween. Note the fact that in this particular geometrythe flat form of the LPD fin implies that the constraint on fin height Hand corrugation width W required to avoid crossover when in the wrappedform may be relaxed. Therefore fins with percent offsets ranging over awider range than is possible when used in the wrapped form may beutilized. It will be understood that flat plate heat exchangers usingthe LPD fin of the subject invention have other structural requirementswhich must be satisfied in order to produce an efficient heat exchanger.For example, for flat plate coolers of a width generally greater than1.5", the inlet and outlet ports must be such to provide rapidtransverse oil flow across the full width of the cooler in order toutilize the full internal area of the heat exchanger. This may beaccomplished in various ways including having transversely elongateinlet and outlet ports extending substantially across the transversewidth of the cooler. Alternatively, the fin may be provided with aregion adjacent the inlet and outlet ports which are specificallystructured to provide rapid transverse flow. Such modifications willgenerally not be required for coolers of width less 1.5".

Similarly, a rectangularly shaped heat exchanger having rounded edgesmay be used instead of a concentric tube heat exchanger with the findimensioned so as to avoid crossover in the corner regions.

In summary, an offset strip fin having a range of dimensions suitablefor cooling of automotive based oils in compact heat exchangers has beendisclosed. The preferred ranges of fin height, corrugation width,thickness and lanced length for a 50% OSF have been determined forautomotive applications of the heat exchanger e.g. using typicaltransmission and transaxle oil at typical oil flow rates in a concentrictube heat exchanger geometry. Fins with offsets different from 50% maybe readily used in the coolers with the fin dimensions being determinedby the geometry of the cooler and wherein studies similar to thosereported above can be carried out to determine the preferred fin heightand corrugation width. Similarly, the heat exchangers and fin structuresof the present invention may be utilized for cooling other liquidsbesides fluids associated with the automotive industry. In this case thepreferred range of lanced lengths can be determined using the liquids tobe cooled in the range of anticipated flow rates.

Therefore, while the present invention has been described andillustrated with respect to the preferred and alternative embodiments,it will be appreciated that numerous variations of these embodiments maybe made without departing from the scope of the invention, which isdefined in the appended claims.

I claim:
 1. An offset strip fin for use in a heat exchanger,comprising:a) a plurality of transverse rows of corrugations, the rowsbeing adjacent and extending in an axial direction, the corrugationshaving a substantially flat top portion and a flat bottom portion, thetop and bottom portions of the corrugations having the same width, thecorrugations having a height in a predetermined range, the corrugationshaving a width in a predetermined range, wherein said height of thecorrugation is greater than said width; and b) the corrugations inadjacent rows of the fin overlapping and interconnected between saidflat top and flat bottom portions, the overlapping corrugations definingperiodically interrupted flow passageways in the axial direction, andwherein the corrugations each have a lanced length in the axialdirection in a predetermined range.
 2. An offset strip fin according toclaim 1 wherein the cross-sectional area of the apertures through thecorrugations in the fluid flow direction is small compared to thesurface area of the corrugations in order to provide short heatconducting paths and a large contact surface area between thecorrugations and the fluid flowing therethrough.
 3. A parallel plateheat exchanger, comprising:a) a generally rectangular metal containerdefining a longitudinal direction, the container having parallel top andbottom plates, means defining an entrance port located adjacent one endof the container and means defining an outlet port located adjacent theopposed end of the container; and b) an offset strip fin disposedbetween the top and bottom plates, the fin being provided with aplurality of transverse rows of corrugations, the rows being adjacentand extending in the longitudinal direction, the corrugations havingflat top portions and flat bottom portions, the top and bottom portionsof the corrugations having the same width, the top portions being inthermal contact with the top plate and the bottom portions being inthermal contact with the bottom plate, each corrugation having parallelside walls, the corrugations having a height in a predetermined range,the corrugations having a width in a predetermined range, wherein saidheight of the corrugations is greater than said width, the corrugationsin adjacent rows of the fin overlapping, the overlapping corrugationsdefining periodically interrupted flow passageways in the longitudinaldirection characterized by laminar fluid flow therethrough, and whereinthe corrugations have a lanced length in the longitudinal direction in apredetermined range suitable to give fully developed periodic flow inthe longitudinal direction.
 4. The heat exchanger according to claim 3wherein the cross-sectional area of the apertures through thecorrugations in the flow direction is small compared to the surface areaof the corrugations in order to provide a short heat conducting path anda large contact surface area between the corrugations and the fluidflowing therethrough.
 5. A heat exchanger according to claim 3 whereinthe lance lengths are in the range suitable to give periodic fullydeveloped flow when the liquid being cooled is flowing therethrough. 6.A heat exchanger according to claim 3 including a transversely elongateinlet port located adjacent one end of the container and a transverselyelongate outlet port located adjacent the opposed end of the container.7. A tubular heat exchanger for cooling transaxle and transmission oil,comprising:a) an outer tube; b) an inner tube disposed within the outertube with the space between the inner tube and the outer tube defining apassageway extending along the axial direction of the tubes; c) an inletport in flow communication with the passageway for admitting fluid to becooled into the passageway; d) an outlet port in flow communication withthe passageway for providing a fluid outlet from the passageway, whereinthe outlet port is spaced from the inlet port; and e) an offset stripfin disposed in the passageway between the inlet and outlet ports,wherein the fin is provided with a plurality of transverse rows ofcorrugations, the rows being adjacent and extending in the axialdirection, the corrugations each having a substantially flat top portionand a flat bottom portion, the top and bottom portions of thecorrugations having the same width, the top portion being in thermalcontact with the inner surface of the outer tube and the bottom portionbeing in thermal contact with the outer surface of the inner tube, thecorrugations having a height in a predetermined range, said corrugationwidth being in a predetermined range, wherein said height of thecorrugations is greater than said width the corrugations in adjacentrows of the fin overlapping and interconnected between said flat top andflat bottom portions, the overlapping corrugations defining periodicallyinterrupted flow passageways in the axial direction, and wherein thecorrugations have a lanced length in the longitudinal direction in apredetermined range.
 8. A heat exchanger according to claim 7 whereinthe inlet port is located adjacent one end of the tubes and the outletport is located adjacent the other end of the tubes.
 9. The tubular heatexchanger according to claim 7 wherein the tubes have a circularcross-section, wherein the passageway between the tubes is an annularpassageway, wherein disposing the fin within the passageway results inthe top portions of adjacent corrugations being separated by a distancegreater than the width of the corrugations and the bottom portions ofadjacent corrugations being separated by a distance which is less thanthe width of the corrugations.
 10. The heat exchanger according to claim9 wherein the fin is fabricated of an alloy from the class of alloyscontaining brass, various steel alloys and various aluminum alloys. 11.The heat exchanger according to claim 10 wherein the fin thickness is inthe range from substantially 0.002" to 0.004".
 12. The heat exchangeraccording to claim 11 wherein the fin height is in the range fromsubstantially 0.100" to 0.130".
 13. The heat exchanger according toclaim 12 wherein the width of the corrugations is in the range fromsubstantially 0.027" to 0.050".
 14. The heat exchanger according toclaim 13 wherein the lanced length is in the range from substantially0.035" to 0.075".
 15. A concentric tube heat exchanger for coolingautomotive transaxle and transmission oil at oil flow rates in the rangefrom substantially 0.50 gpm to 3.5 gpm, comprising:a) an outer tubehaving an inner diameter; b) an inner tube having an outer diameter lessthan the inner diameter of the outer tube, the inner tube beingconcentrically disposed within the outer tube with the space between theinner tube and the outer tube defining an annular passageway extendingalong the axial direction of the tubes, and the concentric tubes beingsealed together at the ends of the tubes; c) the outer and inner tubesdefining an inlet port in flow communication with the annular passagewayfor admitting fluid to be cooled into the passageway, and an outlet portin flow communication with the annular passageway for providing a fluidoutlet from the passageway, wherein the outlet port is spaced from theinlet port; and d) an offset strip fin circumferentially disposed in theannular passageway extending axially between the ends of the tubes,wherein the fin comprises transverse rows of corrugations, thecorrugations defining flow passageways in the axial direction, thecorrugations having a height substantially equal to the differencebetween the inner radius of the outer tube and the outer radius of theinner tube, the corrugations having a flat top portion in thermalcontact with the inner surface of the outer tube and a flat bottomportion in thermal contact with the outer surface of the inner tube, thetop and bottom portions of the corrugations having the same width, thewidth being in a predetermined range, wherein the top portions oftransversely adjacent corrugations are separated by a distance greaterthan the width of the corrugations and the bottom portions oftransversely adjacent corrugations are separated by a distance which isless than the width of the corrugations, and the corrugations having alanced length in the axial direction in a predetermined range.
 16. Aheat exchanger according to claim 15 wherein the inlet port is locatedadjacent one end of the tubes and the outlet port is located adjacentthe other end of the tubes.
 17. The heat exchanger according to claim 15wherein the cross-sectional area of the apertures through thecorrugations in the flow direction is small compared to the surface areaof the corrugations in order to provide a short heat conducting path anda large contact surface area between the corrugations and the fluidflowing therethrough.
 18. The heat exchanger according to claim 15wherein the fin height is in the range from substantially 0.100" to0.130".
 19. The concentric heat exchanger according to claim 18 whereinthe width of the corrugations is in the range from substantially 0.027"to 0.050".
 20. The heat exchanger according to claim 19 wherein thelanced length is in the range from substantially 0.035" to 0.075". 21.The heat exchanger according to claim 20 wherein the fin is fabricatedof an alloy from the class of alloys containing brass, various steelalloys and various aluminum alloys.